Motor or pump assemblies

ABSTRACT

A hydraulic device includes a rotatable drum, at least one piston, a transition arm, and at least one compensating piston. The rotatable drum defines a cylinder that houses the piston. The transition arm is coupled to the piston to translate between linear motion of the piston and rotation of the rotatable drum. The compensating piston is in fluidic communication with the cylinder and is configured to counteract a force caused by an inlet or an output pressure during operation.

This application claims priority under 35 USC § 119(e) to U.S. patent application Ser. No. 60/602,865, filed on Aug. 20, 2004, the entire contents of which are hereby incorporated by reference.

BACKGROUND

This description relates to pumps and motors, for example, hydraulic pumps and motors.

In some hydraulic pumps, rotary motion is translating into reciprocating motion of a piston, which pumps the hydraulic fluid. Conversely, in some hydraulic motors, a pressurized fluid causes reciprocating motion of the piston, which is translated into rotary motion.

SUMMARY

In one aspect, a hydraulic device includes a rotatable drum, at least one piston, a transition arm, and at least one compensating piston. The rotatable drum defines a cylinder that houses the piston. The transition arm is coupled to the piston to translate between linear motion of the piston and rotation of the rotatable drum. The compensating piston is in fluidic communication with the cylinder and is configured to counteract a force caused by an inlet or an output pressure during operation.

Implementations may include one or more of the following features. For example, the hydraulic device may include a shaft and a member, such as a plate, mounted on the shaft. The shaft may be connected to the rotatable drum such that the rotatable drum does not move axially along the shaft. The compensating piston then may be configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member. A flanged nut and a spring also may be mounted to the shaft, with the spring between the member and the flanged nut such that the spring applies a force to the flanged nut in the same direction as the force applied to the member by the compensating piston. The member may be mounted on the shaft such that the member moves axially along the shaft. The compensating piston may be configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member such that the member moves axially along the shaft to compress the spring, resulting in an increase in the force applied to the flanged nut by the spring.

Alternatively, the shaft maybe connected to the rotatable drum such that the rotatable drum moves axially along the shaft. The member, such as a support, may be mounted to the shaft such that the member does not move axially along the shaft. The compensating piston then may be configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member. The compensating piston applying a force to the member may result in a force being applied to the rotatable drum in a direction opposite to the force caused by the inlet or outlet pressure. The hydraulic device may include a U-joint coupled to the member and the transition arm. A spring may be mounted between the member and the rotatable drum such that the spring applies a force to the rotatable drum in the opposite direction of the force caused by the inlet or output pressure during operation.

The hydraulic device may be a hydraulic motor or a hydraulic pump. The hydraulic device may include a bulkhead that defines an inlet or outlet manifold, a cylinder housing the compensating piston, and a port between the inlet or outlet manifold and the cylinder housing the compensating piston. The port may provide the fluidic communication between the compensating piston and the cylinder defined by the rotatable drum. Alternatively, or additionally, the bulkhead may define a port between the cylinder housing the compensating piston and the cylinder defined by the rotatable drum to provide the fluidic communication.

The hydraulic device may include a face valve between the bulkhead and the rotatable drum. The compensating piston may be configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation to maintain a seal between the face valve and the bulkhead or the face valve and the rotatable drum. A spring may be coupled to the compensating piston such that the spring applies a force to the rotatable drum in the opposite direction of the force caused by the inlet or output pressure during operation. The compensating piston may be configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by creating a force that acts on the rotatable drum in a direction opposite to the force caused by the inlet or output pressure.

The hydraulic device may include a non-rotating member coupled to the transition arm. A radial position of the non-rotating member relative to an axis of rotation of the rotatable drum may be adjustable to change an angle between the transition arm and the axis of rotation of the rotatable drum. The transition arm may be coupled to the piston and the non-rotating member such that a change in the angle between the transition arm and the axis of rotation of the rotatable drum changes a stroke of the piston. The non-rotating member may include teeth that mesh with a gear such that rotation of the gear adjusts the radial position of the non-rotating member relative to the axis of rotation of the rotatable drum.

The hydraulic device may include a piston joint assembly coupling the transition arm to the piston. The piston joint assembly may be configured to reduce centrifugal forces acting on the piston as a result of rotation of the rotatable drum and piston. The piston joint assembly may include a casing that defines a hole through which a portion of the piston extends. The hole may have a diameter larger than the portion of the piston extending through the hole. The portion extending through the hole of the casing may terminate in a head piece. The head piece may have a diameter larger than the hole defined by the casing and the casing may be dimensioned to provide a spacing between an inner surface of the casing and the head piece.

In another aspect, a method for counteracting a force in a hydraulic device caused by at least one of an inlet pressure and an outlet pressure during operation may include adjusting a counteracting force as at least one of the inlet pressure or the outlet pressure changes during operation.

In another aspect, a method for counteracting a force in a hydraulic device caused by at least one of an inlet pressure and an outlet pressure may include rotating a drum that defines a cylinder; reciprocating a piston in the cylinder; and communicating a pressure in the cylinder to a compensating piston such that the compensating piston creates a force that counteracts the force caused by at least one of an inlet pressure and an outlet pressure during operation.

Implementations may include one or more advantages. For example, the compensating piston may be able to produce a variable force to counteract the force cause by the inlet or output pressure. The varying force may increase as the inlet or output pressure increases, and decrease as the inlet or output pressure decreases. This may allow the friction between the between the drum and the face valve and between the face valve and bulkhead to be low at low output or inlet pressures, while increasing to maintain a seal between the drum and the face valve and between the face valve and the bulkhead at higher pressures. This may improve efficiency, particularly when the pump or motor is operating at low inlet or output pressures.

In addition, the use of a spring may provide an initial force that is appropriate to counteract the force from the pressure during start-up conditions. The initial force from the spring may be designed so as to maintain the seal between the drum and the face valve and between the face valve and bulkhead during low start-up pressure, without being greater than necessary to maintain the seal by completely or partially counteracting the force generated by the low start-up pressure. In this way, friction during start-up may be decreased, which may improve pump efficiency and may reduce wear.

The details of one or more implementations are set forth in the accompanying drawings and the description below. Other features, objects, and advantages will be apparent from the description and drawings, and from the claims.

DESCRIPTION OF DRAWINGS

FIG. 1A illustrates a combination hydraulic motor/pump.

FIG. 1B is an end view of the combination hydraulic motor/pump taken along line 1B of FIG. 1A.

FIG. 2A and 2B show a face valve of the combination hydraulic motor/pump of FIG 1A and 1B.

FIG. 3 is an end view of the valve cylinder of the combination hydraulic motor/pump of FIGS. 1A and 1B.

FIG. 4 is a side view of the valve cylinder of the combination hydraulic motor/pump of FIGS. 1A and 1B.

FIG. 5A illustrates a hydraulic assembly with reduced centrifugal force effects.

FIG. 5B is a cross-sectional view taken along line 5B of FIG. 5A.

FIG. 5C is a cross-section taken along line 5C in FIG. 5A.

FIG. 5D shows a piston assembly of the hydraulic assembly of FIG. 5A.

FIG. 5E is a cross-sectional view of a piston assembly taken along line 5E of FIG. 5C.

FIG. 5F is a view of the piston assembly and support members taken along line 5F in FIG. 5C.

FIG. 6A illustrates a hydraulic assembly.

FIG. 6B is a cross-sectional view of the hydraulic assembly of FIG. 6A taken along line 6B of FIG. 6A.

FIG. 7A illustrates an alternate implementation of the hydraulic assembly of FIGS. 6A and 6B.

FIG. 7B is a view of a variable pressure mechanism of the hydraulic assembly of FIG. 7A.

FIG. 7C is a view of an alternate implementation of the variable pressure mechanism of the hydraulic assembly of FIG. 7A.

DETAILED DESCRIPTION

Referring to FIGS. 1A and 1B, a combination hydraulic motor/pump 30 particularly useful, e.g., in deep well applications, includes a first side 31 that acts as a hydraulic motor and a second side 32 that acts as a hydraulic pump. The motor side 31 is designed to operate with a high pressure/low volume fluid input and directly drives the pump side, which is designed for higher volume/lower pressure fluid output. Motor/pump 30 is placed within a well or borehole and the space between motor/pump 30 and casing 47 of the well is sealed with a seal 1. Accordingly, when placed below the liquid level of the well or borehole, a low volume/high pressure fluid is pumped into the motor side 31, which then directly drives the pump side 32 to pump the fluid in the well or borehole to the surface at a lower pressure, but higher volume than the fluid pumped into the motor side 31.

Specifically, combination hydraulic motor/pump 30 includes stationary housing parts 2, 9, 13, and 14 that define an open region 35 within motor/pump 30. Located within open region 35 is a transition arm 22. Transition arm 22 is mounted on, e.g., a universal joint (U-joint) 23. For example, U-joint 23 is a cardon type U-joint, however, other types of U-joints can be used. Transition arm 22 includes a nose pin that is coupled to a rotating member 21 via a self-aligning nose pin bearing such that transition arm 21 is at an angle β with respect to assembly axis A. The nose pin is axially fixed within rotating member 21.

Piston assemblies 5 (e.g., three piston assemblies 5) are mounted circumferentially about transition arm 22 via drive pins coupled to piston joint assemblies 6, such as, e.g., one of the piston joint assemblies described in FIGS. 14-16, 23-23A or 56-56F in PCT Application WO 03/100231, filed May 27, 2003, incorporated herein by reference in its entirety, or, e.g., the piston joint assembly described below with respect to FIGS. 5C-5E. Piston assemblies 5 include double-ended pistons having a first piston 7 (the drive piston) on one end and a second piston 3 (the pump piston) on the other end. Pistons 7 are received in cylinders 36 a formed in stationary housing part 9 on motor side 31, while pistons 3 are received in cylinders 36 b formed in stationary housing part 2 on pump side 32. Pistons 7 are centered in cylinder 36 a by seals 8 such that they do not touch the sides of cylinders 36 a. Similarly, pistons 3 are centered in cylinders 36 b by seals 4 such that they do not touch the sides of cylinders 36 b. In an exemplary implementation, the distance between the end face 43 of pistons 3 and end wall 42 of cylinders 36 a when pistons 3 are at the end of their intake stroke (i.e., when pistons 3 are moving in the direction of arrow D) is the same as the distance between the end face 44 of pistons 7 and the end wall 45 of cylinders 36 a when pistons 7 are at the end of their power stroke (i.e., moving in the direction of arrow C). The area of the end face 43 of pistons 3, however, is larger than the area of the end face 44 of pistons 7 (for reasons described further below).

Transition arm 22 is supported on U-joint 23, which is connected to stationary housing part 2 such that U-joint 23 does not rotate. U-joint 23 acts as a pivot point for transition arm 22 such that linear movement of pistons 3 and 7 results in the nose pin of transition arm 22 moving in a generally circular fashion about assembly axis A. Because the nose pin is connected to rotating member 21, the circular movement of the nose pin causes the rotating member 21 to rotate about assembly axis A. Rotating member 21 is connected to a shaft 33, which rotates with rotating member 21.

Referring also to FIGS. 2A, 2B, 3, and 4, mounted on shaft 33 is a face valve 10 and a valve cylinder 11 that control the flow of fluid in motor half 31. Face valve 10 and valve cylinder 11 are mounted on shaft 33 such that face valve 10 and valve cylinder 11 rotate with shaft 33. For instance, shaft 33 has a splined end that mates with splined holes 37 and 40 defined in face valve 10 and valve cylinder 11, respectively.

Shaft 33 is attached to valve cylinder 11 via a bolt 18. Between the bolt head and valve cylinder 11 is a spring washer 19, such as a Belleville washer. Spring washer 19 provides a force against valve cylinder 11 so that valve cylinder 11 and face plate 10 remain in contact during operation, thereby maintaining a seal between the two to limit the possibility of leakage of the drive fluid.

Face valve 10 includes an inlet section 38 that is connected to an inlet tube 16 through a port 12 and space 41 in valve cylinder 11. Inlet tube 16 extends from space 41 in valve cylinder 11 through stationary housing part 14. Valve cylinder 11 rotates about inlet tube 16 during operation and o-rings 17 seal the area between valve cylinder 11 and inlet tube 16. Face valve 10 also includes an outlet section 39 connected to a port 20 of valve cylinder 11. Port 20 is connected to holes 34 located about stationary housing part 13 such that the drive fluid is exhausted from motor side 31 into the well.

During operation, fluid is delivered from the surface via a low volume/high pressure line to inlet tube 16 and delivered to inlet section 38 via port 12 in valve cylinder 11. Inlet section 38 is in fluidic contact with some of cylinders 36 a (e.g., two of the cylinders 36 a) at a time as valve cylinder 11 and face valve 10 rotate. Thus, the fluid from the surface enters the aligned cylinders 36 a and causes the corresponding pistons 7 to move along piston axis P in the direction of arrow C (FIG. 1A). The movement of the aligned pistons 7 results in movement of the nose pin of transition arm 22 in a generally circular fashion about assembly axis A, which cause rotating member 21 and shaft 33 to rotate. As a result, face valve 10 and valve cylinder 11 also rotate about assembly axis A, which brings other cylinders 36 a and corresponding pistons 7 into alignment with inlet section 38, thereby allowing the delivered fluid to cause those pistons 136 a to move along piston axis P in the direction of arrow C.

For simplicity of illustration, port 12 is shown as aligned with cylinder 36 a in FIG. 1B while drive piston 7 is at the end of its power stroke and getting ready to enter an exhaust stroke. During operation, however, valve cylinder 11 is rotated 90° from the position shown when the stroke of drive piston 7 is at this point. That is, cylinder 36 a is aligned with one of the ends 46 a or 46 b (depending on the direction of rotation) of inlet section 38.

At the same time that some of the pistons 7 are moving along piston axis P in the direction of arrow C, the piston(s) 7 not aligned with inlet section 38 are moving along piston axis P in the opposite direction, i.e., in the direction of arrow D. The cylinders 36 a corresponding to the piston(s) 7 moving in the direction of arrow D are aligned with outlet section 39. Thus, as the pistons 7 aligned with outlet section 39 move in the direction of arrow D, fluid contained in their corresponding cylinders 36 a (which entered those cylinders through inlet section 38) is expelled through outlet section 39 to port 20 and out of motor/pump 30 through holes 34.

In this manner, motor half 31 acts as a hydraulic motor. Delivery of the drive fluid to inlet tube 16 causes the pistons 7 to reciprocate along piston axis P. The reciprocation of pistons 7 results in pistons 3 also reciprocating along piston axis P in cylinders 36 b.

Cylinders 36 b are in fluidic communication with inlet ports 28 located on pump half 32. Inlet ports 28 open to one side of seal 1 such that they are in fluidic communication with the fluid to be pumped. Inlet ports 28 include a check valve 27 (e.g. a poppet valve) that only allows fluid to flow through inlet ports 28 in the direction of arrow D. Consequently, as pistons 3 move in the direction of arrow D, fluid is pulled into inlet ports 28 and cylinders 36 b; however, as pistons 3 move in the direction of arrow C, fluid does not flow out of pump half 32 through inlet ports 28.

Instead, the fluid in cylinders 36 b are output via output ports 24, which are in fluidic communication with inlet ports 28 via a port 26. Outlet ports 24 are open to the opposite side of seal 1 than inlet ports 28 such that the fluid is pumped towards the top of the well or borehole. The fluid is guided towards the top of the well or borehole by the well casing 47. Outlet ports 24 also include a check valve 25 that only allows fluid to flow in the direction of arrow D. Consequently, as pistons 3 move in the direction of arrow C, fluid is pumped out of cylinders 36 b, through ports 26, and out of outlet ports 24; however, as pistons 3 move in the direction of arrow D, fluid is not pulled into pump half 32 through outlet ports 24.

Accordingly, fluid from the surface entering cylinders 36 a that are aligned with inlet section 38 of face valve 10 causes the corresponding pistons 7 and 3 to move in the direction of arrow C. This results in fluid being pumped out of the corresponding cylinders 36 b through ports 26 and out outlet ports 24. At the same time, the pistons 7 on the motor half 31 not aligned with inlet section 38 and the corresponding pistons 3 on the pump side 32 are moving in the direction of arrow D. This results in the fluid in the corresponding cylinders 36 a being pumped into outlet section 39 of face valve 10, through outlet port 20 of valve cylinder 11, and out of holes 34, along with fluid in the well or borehole being pulled into the corresponding cylinders 36 b on the pump side 32 via inlet ports 28.

Thus, during operation, a high pressure/low volume down line is connected to the inlet port 16 on motor side 31, and combination hydraulic motor/pump 30 is placed below the liquid level in the well or borehole such that inlet ports 28 on pump side 32 are in fluidic communication with the fluid in the well or borehole to be pumped (the pump fluid). A high pressure/low volume drive fluid is pumped to the inlet tube 16 on motor side 31 via the down line. The drive fluid enters some of cylinders 36 a by port 12 of valve cylinder 11 and inlet section 38 of face valve 10 and cause pistons 7 to move linearly within cylinders 36 a, which causes rotating member 21 and, consequently, face valve 10 and valve cylinder 11 to rotate. As face valve 10 and valve cylinder 11 rotate, the drive fluid causes other of pistons 7 to move linearly within cylinders 36 a, and pistons 136 a begin reciprocating in cylinders 36 a. This expels the drive fluid in cylinders 36 a out through outlet section 39 in face valve 10, port 20 in valve cylinder 11, and holes 34.

The movements of pistons 7 also directly drive pistons 3, thereby causing pistons 3 to reciprocate. The reciprocating motion of pistons 3 pumps the fluid in the well or borehole into cylinders 36 b through inlet ports 28, and out ports 26 to outlet ports 24. Because the area of the end faces 43 of pistons 3 is greater than the area of end faces 44 of pistons 7, pumped fluid at a lower pressure than the drive fluid, but a higher volume than the drive fluid, is pumped back up. The ratio of the area of the end faces 43 of pistons 3 to the area of end faces 44 of pistons 7 determines the ratio of the volume pumped to the volume delivered and can be, e.g., in the range of 3:1 to 10:1.

Ideally, the ratio of volume pump to volume delivered would be exactly equal to the ratio of the area of the end faces 43 of pistons 3 to the area of end faces 44 of pistons 7. Absent friction and with an incompressible fluid, the product of the pressure times fluid capacity of the drive fluid would be the same as the pumped fluid. The combined area of end faces 43, in FIGS. 1A and 1B, is three times that of the combined area of end faces 44. Thus, the volume of drive fluid supplied to motor part 31 through inlet port 16 ideally produces three times the volume of pump fluid out of ports 24 at one-third the pressure. For example, if the drive fluid is supplied at a rate and pressure of 10 gallons per minute and 9,000 PSI respectively, the combined output of output ports 24 is ideally 30 gallons per minute at 3,000 PSI.

But, since the friction and the volumetric efficiency is unlikely to be zero, this ideal is unlikely to be achieved. However, the efficiency of hydraulic motor/pump 30 can be in the mid-ninety percent range. This is because the force provided by pistons 7 is transferred directly to pistons 3, without passing through the rotating parts of hydraulic motor/pump 30, which instead serve simply to control the timing of pistons 7 and 3 so that they are evenly spaced in time for proper motoring and pumping purposes. In addition, pistons 7 and 3 are moving in straight lines and generate little or no side load on cylinders 36 a and 36 b. Consequently, the frictional losses are low.

Accordingly, for example, if the ratio of the area of end faces 43 to the area of end faces 44 is 3:1 and efficiency is ninety percent, then for every gallon of drive fluid delivered to motor half 110 a, approximately 2.7 gallons of pump fluid would be returned (in addition to the one gallon of drive fluid).

Combination hydraulic motor/pump 30 advantageously allows the drive fluid to be the same or compatible with the pumped fluid, such that a return hydraulic line is not needed. Rather, both fluids are returned up the well or borehole. For instance, if potable water is being pumped, then high pressure water can be used as the drive fluid. Similarly, if the fluid being pumped is crude oil, then a fluid such as cleaned up crude oil, semi-refined oil, or even water can be used to drive the motor so that the drive fluid and pumped fluid are able to be returned along the same up line. In the event an application requires that the drive fluid and pumped fluid not mix, the combination hydraulic motor/pump 30 can be implemented with a hydraulic return line for the drive fluid.

The fluid pressure needed to run combination hydraulic motor/pump 30 is generated at the top of the well, for example, by an electric motor or internal combustion engine used to run a hydraulic pump that provides a high pressure fluid flow. The ratio of fluid pressure into the well to the fluid pressure out of the well is roughly the inverse of the respective flow rates into and out of the well, with allowances for losses in the pipes owing to viscosity and other mechanical forces. Because the fluid/power source is located at the top of the well or borehole, it is easily serviced, resulting in low maintenance costs. The power source may be a combination internal combustion engine/hydraulic pump similar to combination hydraulic motor/pump 30 (with motor side 31 implemented as an internal combustion engine, such as is described with respect to FIGS. 32 and 32a in PCT Application WO 03/100231, filed May 27), which provides high energy efficiency and low power costs. Similarly, a combination electric motor/hydraulic pump similar to that described with respect to FIGS. 65 and 67 in PCT Application WO 03/100231, filed May 27, 2003, can be employed.

Changing the pressure or rate of flow of the drive fluid in the high pressure drive line changes the rate of flow of the pumped fluid in the low pressure return. When a combination engine and hydraulic pump are used as the power source, changing the engine speed varies the rate of flow of the drive fluid. Alternatively, the hydraulic pump portion of the power source can be a variable stroke pump (such as the ones described with respect to FIG. 50 and 54 in PCT Application WO 03/100231, or described below with respect to FIGS. 5A and 5B), which allows the flow rate to be changed without varying the engine speed. This provides a simple mechanism for controlling the rate of pumping.

A number of implementations have been described. Nevertheless, it will be understood that various modifications may be made. For example, the flow of fluid can be controlled on pump side 32 through a second face valve whose rotation is synchronized with face valve 10, instead of through the check valves 28 and 25. In addition, a variable stroke mechanism can be included in combined motor/pump 30, such as the ones described with respect to FIG. 50 and 54 in PCT Application WO 03/100231, or described below with respect to FIGS. 5A and 5B. Furthermore, while three double ended pistons and corresponding cylinders have been shown, more or less pistons and cylinders can be used.

Referring to FIGS. 5A-5C, a hydraulic assembly 500 (e.g., hydraulic pump or motor) designed to reduce misalignments caused by centrifugal forces includes a stationary housing 502 defining a chamber 504, and a rotating drum 506 located within chamber 504. Drum 506 is mounted on a shaft 508. Application of torque to shaft 508 rotates drum 506 about assembly axis A in housing 502, and rotation of drum 506 causes rotation of shaft 508. Drum 506 includes a first section 506 a, a second section 506 b, and support members (e.g., cylindrical tie rods) 506 c that connect first section 506 a and second section 506 b so that the first and second sections rotate together (and for other reasons further described below).

First section 506 a and second section 506 b define an open region 512 between them. Located within open region 512 is a transition arm 514. Transition arm 514 is mounted to, e.g., a universal joint (U-joint) 516. For example, U-joint 516 is a cardon type U-joint, however, other types of U-joints can be used. Transition arm 514 includes a nose pin 518 that is coupled to a bearing block 520 via a self-aligning nose pin bearing 558 such that transition arm 514 is at an angle β with respect to assembly axis A. Nose pin 518 is axially fixed within bearing block 520.

Bearing block 520 is received in an arced channel 560 defined in housing 502 such that bearing block does not rotate, and includes a gear-toothed surface 562 that mates with a pinion gear 564 within housing 502. A mechanism (not shown), such as an external knob with a shaft engaged with pinion gear 564, is used to turn pinion gear 564. Turning pinion gear 564 causes bearing block 520 to slide in arced channel 560 to and away from assembly axis A, thereby changing the angle β. Changing the angle β changes the piston stroke and, consequently, the motor/pump capacity. This allows an operator to adjust the piston stroke of assembly 500. In other implementations, other mechanisms of controlling the angle β are employed, such as the mechanisms described with respect to FIG. 50 and 54 in PCT Application WO 03/100231, filed May 27, 2003.

Transition arm 514 also includes drive pins 522 coupled to piston assemblies 524 (e.g., seven piston assemblies 524) via piston joint assemblies 526, such as, e.g., piston joint assemblies described below with respect to FIGS. 5C-5E or, e.g., the piston joint assemblies described in FIGS. 56-56F in PCT Application WO 03/100231, filed May 27, 2003. Piston assemblies 524 include single ended pistons having a piston 556 on one end. Pistons 556 are received in cylinders 530 formed in first section 506 a.

During operation as a hydraulic pump, shaft 508 is rotated, causing drum 506 to rotate, which results in pistons 556 and transition arm 514 rotating with drum 506. Because nose pin 518 is fixed and transition arm 514 is at an angle β with respect to assembly axis A, rotation of drum 506 causes pistons 556 to reciprocate in cylinders 530 along piston axis P. As pistons 556 reciprocate, fluid is pumped from an inlet 582 to an outlet 538. Inlet 582 connects to an inlet manifold 584, while outlet 538 connects to an outlet manifold 570.

Pump/motor 500 includes a face valve 536 to control the fluid pumping during operation. Face valve 536 allows fluid to be pulled from the inlet 582 through inlet manifold 584 into a cylinder 530 when the cylinder's corresponding piston 556 is on an intake stroke (moving in the direction of arrow F), while directing fluid from a cylinder 530 through outlet manifold 570 to outlet 538 when the cylinder's corresponding piston 556 is on a pump stroke (moving in the direction of arrow G).

During operation as a hydraulic motor, the operation is reversed. A pressurized stream of fluid is provided to inlet 582. This fluid causes some of the pistons 556 to move in the direction of arrow F. This movement causes drum 506 and transition arm to 514 to rotate, which causes the other pistons 556 to move in the direction of arrow G. The movement of these pistons expels any drive fluid in their corresponding cylinders 530 out of outlet 538. As this process continues, pistons 556 reciprocate in cylinders 530 as drum 506 rotates. Rotation of drum 506 causes shaft 508 to rotate.

Referring to FIGS. 5D-5F, piston joint assemblies 526 include a spherical or cylindrical bearing 540 that is coupled to drive pin 522. Bearing 540 is seated between bearing pads 542 a and 542 b located in a casing 544. Casing 544 includes a piston end 544 a and an end 544 b opposite the piston end. A hole 546 is formed in piston end 544 a. In addition, piston end 544 a and bearing pad 542 a define a space 534 therebetween.

Piston 556 has a circular head 548 that is received in space 534 and abuts against a face of bearing pad 542 a, while piston 556 projects through hole 546. Circular head 548 has a diameter larger than hole 546 such that circular head 548 can not pass through hole 546. Hole 546 has a diameter larger than the outer diameter of piston 556 passing through hole 546. In addition, there is spacing 554 between the inner surface of the sides of casing 544 and circular head 548. The larger diameter of hole 546 and the spacing 554 between the inner surface of casing 544 and circular head 548 allows casing 544, bearing 540, and bearing pads 542 a and 542 b to slide relative to circular head 548 and piston 556 in the direction of arrow E for reasons described further below. The faces of circular head 548 are polished to minimize friction.

Referring to FIGS. 5B, 5E, and 5F, each side 544 a, 544 b of casing 544 includes two extensions 550, each having, e.g. half circle, cutouts 552 that mate with support members 506 c on either side of the piston joint assemblies 526, as shown in FIG. 5B, so that piston joint assemblies 526 ride along support members 506 c. Support members 506 c are polished and extensions 550 have a low friction surface to reduce friction as piston joint assemblies 526 ride along support members 506 c.

Support members 506 c, extensions 550 with cutouts 552, and the sliding of piston 556 relative to casing 544 act to reduce the effects of centrifugal force. During operation as a pump, the rotation of drum 506 and piston assemblies 524 results in transition arm 514 (through drive pins 522) driving piston joint assemblies 526 back and forth along piston axis P (i.e., causing piston joint assemblies 526 to reciprocate along piston axis P). Piston joint assembly 526 applies this drive force to piston 556 through circular head 548. On the pump stroke, this force is applied to circular head 548 by bearing pad 542 a. On the intake stroke, this force is applied to circular head 548 by the inner surface of piston end 544 a of case 544. Similar forces are applied during operation as a motor.

The rotation of drum 506 and piston assemblies 524 produces centrifugal force on piston joint assemblies 526, causing them to move outward (i.e., radially away from assembly axis A). This movement of piston joint assemblies 524 can cause misalignment of pistons 556 in cylinders 530 if the movement also acts on pistons 556. The attachment of piston joint assemblies 524 to support members 506 c via cutouts 552 reduces the magnitude of this movement by transferring the centrifugal force of the piston joint assemblies 526 to support members 506 c. In addition, the ability of piston 556 to slide relative to piston joint assembly 526 isolates piston 556 from any outward movement of piston joint assembly 526 that does occur. Thus, any movement of piston joint assembly 526 that does occur is prevented from being applied to pistons 556, particularly if the diameter of hole 546 and spacing 554 are designed such that circular head 548 and piston 556 can slide without contacting casing 544 for the expected amount of movement of piston joint assembly 526.

As a result, the centrifugal force acting on pistons 556 is reduced to the centrifugal force generated by the mass of pistons 556, rather than the centrifugal force generated by the mass of pistons 556 and piston joint assemblies 526. The mass of pistons 556 can be kept relatively small, in the range of one-sixth to one-eighth of the mass of piston joint assemblies 526. Consequently, pistons 556 are able to have a closer fit to their respective cylinders 530 because the frictional heating and resulting size change is lower than when the centrifugal force acting on pistons 556 is the result of the mass of pistons 556 and piston joint assemblies 526. When the centrifugal force acting on pistons 556 is the result of the mass of pistons 556 and piston joint assemblies 526, pistons 556 experience greater misalignment and, therefore, rubbing against the cylinder sidewalls. This results in frictional heating that increases the size of pistons 556, which causes a greater surface area of pistons 556 to touch the cylinder sidewalls if not enough spacing is providing between pistons and the cylinder sidewalls.

A number of implementations have been described. Nevertheless, it will be understood that various modifications may be made. For example, other mechanisms to control fluid flow can be used; for example, poppet valves such as those shown in FIG. 1 can be used. Support members 506 c, extensions 550 with cutouts 552, and the sliding of piston 556 relative to casing 544 can be implemented in other types of assemblies that use rotating pistons and cylinders. In addition, while seven single ended pistons 556 and corresponding cylinders 530 have been shown, more or less pistons and cylinders can be used and the pistons can be double-ended, rather than single ended, as shown, for example, in the assembly described with respect to FIG. 1.

Referring to FIGS. 6A and 6B, a hydraulic assembly 600 (e.g., hydraulic pump or motor) designed to counteract forces that tend to move the face valve 636 away from the rotating drum 606 or bulkhead 684, similar to pump/motor 500, includes a stationary housing 602 defining a chamber 604, and a rotating drum 606 located within chamber 604. Drum 606 is mounted on a shaft 608 such that drum 606 does not slide along shaft 608. For example, shaft 608 fits through a hole in drum 608 and a pin is used to prevent sliding and causes drum 606 to rotate with shaft 608. Alternatively, a hole in drum 606 can be press fit to shaft 608. Drum 606 includes a first section 606 a, a second section 606 b, and connecting members 606 c that connect first section 606 a and second section 606 b so that the first and second section rotate together.

First section 606 a and second section 606 b define an open region 612 between them. Located within open region 612 is a transition arm 614. Transition arm 614 is mounted to, e.g., a universal joint (U-joint) 616. For example, U-joint 616 is a cardon type U-joint, however, other types of U-joints can be used. Transition arm 614 includes a nose pin 618 that is coupled to a bearing block 620 via a self-aligning nose pin bearing 658 such that transition arm 614 is at an angle β with respect to assembly axis A. Nose pin 618 is axially fixed within bearing block 620. Bearing block 620 is mounted in an arced channel 660 such that bearing block 620 does not rotate and allows the angle β to be adjusted as described above with respect to FIG. 5A.

Transition arm 614 also includes drive pins 622 coupled to piston assemblies 624 (e.g., seven piston assemblies 624) via piston joint assemblies 626, such as, e.g., piston joint assemblies described above with respect to FIGS. 5D-5F (and assembly 600 may include support members 506 c) or, e.g., the piston joint assemblies described in FIGS. 56-56F in PCT Application WO 03/100231, filed May 27, 2003. Piston assemblies 624 include single ended pistons having a piston 656 on one end and a guide rod 628 on the other end. Pistons 656 are received in cylinders 630 formed in first section 606 a. Guide rods 632 are received in a sleeve-bearing 688 held by second section 606 b.

During operation as a pump, as shaft 608 is rotated, drum 606 rotates, which results in pistons 656 and transition arm 614 rotating with drum 606. Because nose pin 618 is fixed and transition arm 614 is at an angle β with respect to assembly axis A, rotation of drum 606 causes pistons 656 to reciprocate in cylinders 630 along piston axis P. As pistons 656 reciprocate, fluid is pumped from an inlet (not shown, but similar to inlet 582) to an outlet 638. Pump/motor 600 includes a face valve 636 (like face valve 10) to control the fluid pumping during operation. Face valve 636 allows fluid to be pulled from the inlet through an inlet manifold (not shown, but similar to inlet manifold 584) when the cylinder's corresponding piston 656 is on an intake stroke (moving in the direction of arrow F), while directing fluid from a cylinder 630 to through outlet manifold 670 to outlet 638 when the cylinder's corresponding piston 656 is on a pump stroke (moving in the direction of arrow G).

During operation as a hydraulic motor, the operation is reversed. A pressurized stream of fluid is provided to the inlet. This fluid causes some of the pistons 656 to move in the direction of arrow F. This movement causes drum 606 and transition arm to 614 to rotate, which causes the other pistons 656 to move in the direction of arrow G. The movement of these pistons expels any drive fluid in their corresponding cylinders 630 out of outlet 638. As this process continues, pistons 656 reciprocate in cylinders 630 as drum 606 rotates. Rotation of drum 606 causes shaft 608 to rotate.

The pressure of the hydraulic fluid through face valve 636, whether from the inlet when operated as a motor or from outlet 638 and pistons 656 when operated as a pump, causes face valve 636 to separate from drum 606 or bulkhead 684, which can result in leakage and valve failure. To counteract this force, a variable force mechanism 682 is included. The variable force mechanism 682 includes compensating pistons 662 (e.g., two compensating pistons 662) located in cylinders 664 defined by bulkhead 684, a plate 666, a washer spring 668, and a nut 678 with flange 680. Compensating pistons 662 and cylinders 664 are situated parallel to and on opposite sides of assembly axis A. Cylinders 664 are connected via ports 672 to the inlet or output manifold 670 that leads from face valve 636 to outlet 638. When assembly 600 is operated as a pump, cylinders 664 are connected to outlet manifold 670. Conversely, when assembly 600 is operated as a motor, cylinders 664 are connected to the inlet manifold.

Heads 674 of compensating pistons 662 contact plate 666. Plate 666 is mounted around shaft 608 such that plate 666 does not rotate when shaft 608 rotates, but plate 666 can slide linearly along assembly axis A. For example, the outer diameter of plate 666 may be sized such that it has a close fit with the inner diameter of casing 602 in the space 686 so that plate 666 is supported and can slide linearly along assembly axis A. Plate 666 then has a hole (not shown) through which shaft 608 passes. The hole's diameter is a sufficient amount so that shaft 608 may pass through a hole in plate 666 without touching plate 666. A tang (not shown) extends from plate 666 and fits into a groove (not shown) in the inner wall of casing 602 in space 686 to prevent plate 666 from spinning.

Also mounted on shaft 608 next to plate 666 is a thrust bearing 676, spring washer 668, and nut 678 with flange 680. Nut 678 and spring washer 668 are situated on shaft 608 such that they rotate with shaft 608, but do not slide along shaft 608. For example, a portion of shaft 608 is threaded. After spring washer 668 is placed onto shaft 608, nut 678 is screwed onto the threaded portion. Nut 678 is screwed onto shaft 608 to a position that partially compresses spring 668, thereby exerting a force in the direction of arrow G on nut 678 through flange 680.

In addition, as compensating pistons 662 exert a force on plate 666 (as described below), plate 666 moves in the direction of arrow G. Movement of plate 666 applies the force exerted on plate 666 to spring washer 668 through thrust bearing 676, which results in spring washer 668 being further compressed, thereby increasing the force exerted in the direction of arrow G on nut 676 through flange 680. Because drum 606 is mounted on shaft 608 such that drum 606 does not slide along shaft 608, and nut 678 does not slide along shaft 608, the force from spring 668 on nut 678 is exerted on drum 606 through shaft 608, thereby urging drum 606 towards face valve 636. This force acts to counteract the force in the direction of arrow F caused by the hydraulic pressure.

When there is no inlet or output pressure (e.g., before pump/motor 600 begins operation), spring 668 provides an initial force on drum 606 in the direction of arrow G because spring 668 is partially compressed as described above. This initial force maintains a seal between drum 606 and face valve 636, and between face valve 636 and bulkhead 684, thereby preventing the leakage of fluid when pump/motor 600 begins operating. The initial force exerted by spring 668 depends on the amount spring 668 is initially compressed, which depends on the position nut 678 is mounted on shaft 608. The amount of the initial force is sufficient to maintain a seal between drum 606 and face valve 636 and between face valve 636 and bulkhead 684 during start-up conditions.

During operation, a force proportional to the input pressure, when operated as a motor, or output pressure, when operated as a pump, is exerted on drum 606 in the direction of arrow F. This force tends to urge face valve 636 away from bulkhead 684 in the case of a motor, and drum 606 away from face valve 636 in the case of a pump. To compensate for this force, the pressure is communicated from inlet (for a motor) or output manifold 670 (for a pump) to compensating pistons 662 by ports 672. The pressure causes compensating pistons 662 to move in the direction of arrow G, thereby exerting a force on plate 666, which is transferred to drum 606 through bearing 676, spring 668, nut 678, and shaft 608. This compensating force counteracts the force acting on drum 606 and/or face valve 636 in the direction of arrow F so as to maintain the seal between drum 606 and face valve 636, and between face valve 636 and bulkhead 684.

The amount of compensating force exerted by one of the compensating pistons 662 on plate 666 (and hence drum 606) is proportional to the input or output pressure times the area of the end face of compensating piston 662. Thus, as the input or output pressure increases, the force exerted on drum 606 in the direction of arrow F increases, and the compensating force exerted by compensating pistons 662 in the direction of arrow G also increases. The total compensating force exerted on drum 606 depends on the total compensating force exerted by compensating pistons 662 and the spring force of spring 668. Accordingly, spring 668, the number of compensating pistons 662, and the area of the compensating pistons 662 are designed to exert an additional force for a given pressure that, when added to the initial force provided by spring 668, is sufficient to counteract the force in direction F that results during operation.

The use of spring 668 to provide an initial force and compensating pistons 662 to provide additional compensating force proportional to the input or output pressure allows the friction between drum 606 and face valve 636 and between face valve 636 and bulkhead 684 to be low at low inlet or output pressures, while increasing to maintain a seal between drum 606 and face valve 636 and between face valve 636 and bulkhead 684 at higher pressures. This improves efficiency when the pump or motor is operating at low input or output pressures and also reduces the friction and wear during start-up conditions.

Referring to FIGS. 7A and 7B, an alternate implementation of a hydraulic assembly 700 (e.g., hydraulic pump or motor) designed to counteract forces that tend to move the face valve 736 away from the rotating drum 706 or bulkhead 784, which is designed and operated similar to pump/motor 600, includes a variable force mechanism 782 having compensating pistons 762 located in cylinders 764 defined in drum 706. Cylinders 764 are positioned parallel to cylinders 730 and circumferentially spaced about assembly axis A. Cylinders 764 are fluidically connected to cylinders 730 (and, hence, the input or output pressure) by ports 772. There are, for example, the same number of compensating pistons 762 as there are pistons 756. However, the same number of compensating pistons 762 as pistons 756 does not need to be used. For instance, for an even number of pistons 756, half the number of compensating pistons 762 can be used (e.g., 4 compensating pistons 762 for 8 pistons 756).

A first end portion 774 of pistons 762 contact a support portion 780 of U-joint 716, which is mounted to shaft 708 such that it rotates with, but does not slide along shaft 708. Furthermore, because U-joint 716 is connected to transition arm 714, which is connected to bearing block 720 located in an arced channel (not shown in FIG. 7), when pistons 762 exert a force on support portion 780, U-joint 716 does not move linearly along assembly axis A.

In addition, drum 706 is mounted to shaft 708 such that drum 706 can slide along shaft 708 and rotate with shaft 708. For example, shaft 708 has a splined end that mates with a splined hole in drum 706. A spring washer 768 (e.g., a Belleville washer) is mounted on shaft 708 between drum 706 and support portion 780. Spring washer 768 is partially compressed.

As with pump/motor 600, when there is no input or output pressure (e.g., before pump/motor 700 begins operation), spring 768 provides an initial force on drum 706 in the direction of arrow G because spring 768 is partially compressed as described above. This initial force maintains a seal between drum 706 and face valve 736 and between face valve 736 and bulkhead 784, thereby preventing the leakage of fluid when pump/motor 700 begins operating. The initial force exerted by spring 768 depends on the amount spring 768 is initially compressed.

During operation, as the pistons are reciprocating, a force proportional to the inlet or output pressure is exerted on drum 706 in the direction of arrow F, which is compensated for by compensating pistons 762. The inlet or output pressure is communicated from cylinders 730 to compensating pistons 762 by ports 772. The pressure in cylinders 730 causes a force to be exerted at a second end 774 b of compensating piston 762, which results in compensating piston 762 applying the force to support portion 780, causing an equal but opposite force to be exerted on an end wall 764 a of cylinder 764 in the direction of arrow G. Because support portion 780 does not move linearly along assembly axis A, but drum 706 does slide along shaft 708, the forces cause drum 706 to be urged in the direction of arrow G. The force acting on drum 706 from the pressure counteracts the force acting on drum 706 in the direction of arrow F so as to maintain the seal between drum 706 and face valve 736.

The amount of compensating force exerted on drum 706 is proportional to the inlet or output pressure times the cross-sectional area of the compensating piston 762. Thus, as the inlet or output pressure increases and, consequently, the force exerted on drum 706 in the direction of arrow F increases, the compensating force exerted in the direction of arrow G increases. The total compensating force exerted on drum 706 thus depends on the cross-sectional area of the pistons 762 and the number of pistons 762. Accordingly, the number of compensating pistons 762 and the cross-sectional area of the compensating pistons 762 are designed to exert an additional force for a given pressure that, when added to the initial force provided by spring 768, is sufficient to counteract the force in direction F that results during operation.

As with pump/motor 600, the use of spring 768 to provide an initial force and compensating pistons 762 to provide additional compensating force proportional to the input or output pressure allows the friction between drum 706 and face valve 736 to be low at low input or output pressures, while increasing to maintain a seal between drum 706 and face valve 736 at higher pressures. This improves efficiency when the pump is operating at low pressures and also reduces the friction and wear during start-up conditions. In addition, the design of pump/motor 700 allows pump/motor 700 to be more compact in design than pump/motor 600.

Referring to FIG. 7C, in an alternate implementation of the assembly of FIGS. 7A and 7B, rather than spring washers 768, an assembly 700 includes springs 784, such as coil springs, coupled to the compensating pistons 762, to provide an initial compensating force. Compensating piston 762 includes a pin 762 a at end portion 774 b near port 772. Pin 762 a has a smaller diameter than the rest of piston 762. Coil spring 784 is seated in cylinder 764 between an end face 764 a of cylinder 764 and an end face portion 774d of piston 762, with pin 762 a of received within coil spring 784 to center coil spring 784 in cylinder 764.

Each of the coil springs 784 are partially compressed, which causes springs 784 to exert a force on end face portion 774d of compensating pistons 762 in the direction of arrow F and an equal but opposite force on an end wall 764 a of cylinder 764 in the direction of arrow G. Because support portion 780 does not move linearly along assembly axis A, but drum 706 does slide along shaft 708, the forces cause drum 706 to be urged in the direction of arrow G. The compensating force exerted on drum 706 in direction G by springs 784 is sufficient to maintain a seal between drum 706 and face valve 736 and between face valve 736 and bulkhead 784 during start-up conditions, thereby preventing the leakage of fluid when pump/motor 700 begins operating. Then, as described above with respect to FIGS. 7A and 7B, during operation, the compensating pistons 762 result in an additional force that depends on the input or output pressure, which maintains the seal between drum 706 and face valve 736 and between face valve 736 and bulkhead 784.

A number of implementations have been described. Nevertheless, it will be understood that various modifications may be made. For example, other mechanisms to control fluid flow can be used, such as poppet valves as described with respect to FIG. 1. In addition, other numbers of pistons and cylinders can be used, and the pistons can be double ended, rather than single ended, as shown in FIG. 1.

Furthermore, elements of one or more implementations described above may be combined, deleted, supplemented, or modified to form further implementations. For example, the variable force mechanisms of FIGS. 6A-6B or 7A-7B may be incorporated into assembly 500, and supports 506 c and piston joint assemblies 526 may be used in assemblies 600 or 700. Accordingly, other implementations are within the scope of the following claims. 

1. A hydraulic device comprising: a rotatable drum defining at least one cylinder; at least one piston housed in the cylinder; a transition arm coupled to the at least one piston and configured to translate between linear motion of the piston and rotation of the rotatable drum; and at least one compensating piston in fluidic communication with the cylinder and configured to counteract a force caused by at least one of an inlet pressure and an output pressure during operation.
 2. The hydraulic device of claim 1 further comprising: a shaft connected to the rotatable drum such that the rotatable drum does not move axially along the shaft; a member mounted on the shaft; wherein the at least one compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member.
 3. The hydraulic device of claim 2 further comprising: a flanged nut mounted to the shaft; a spring mounted to the shaft between the member and the flanged nut such that the spring applies a force to the flanged nut in the same direction as the force applied to the member by the compensating piston.
 4. The hydraulic device of claim 3 wherein the member is mounted on the shaft such that the member moves axially along the shaft and the at least one compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member such that the member moves axially along the shaft to compress the spring, resulting in an increase in the force applied to the flanged nut by the spring.
 5. The hydraulic device of claim 2 wherein the member comprises a plate.
 6. The hydraulic device of claim 1 comprising a hydraulic motor.
 7. The hydraulic device of claim 6 further comprising: a bulkhead defining: an inlet manifold connected to an inlet; a cylinder housing the compensating piston; and a port between the inlet manifold and the cylinder housing the compensating piston to provide the fluidic communication between the compensating piston and the cylinder defined by the rotatable drum.
 8. The hydraulic device of claim 1 comprising a hydraulic pump.
 9. The hydraulic device of claim 8 further comprising: a bulkhead defining: an outlet manifold connected to an outlet; a cylinder housing the compensating piston; and a port between the cylinder housing the compensating piston and the outlet manifold to provide the fluidic communication between the compensating piston and the cylinder defined by the rotatable drum.
 10. The hydraulic device of claim 1 further comprising: a shaft connected to the rotatable drum such that the rotatable drum moves axially along the shaft; a member mounted to the shaft such that the member does not move axially along the shaft; and wherein the at least one compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member.
 11. The hydraulic device of claim 10 further comprising a U-joint coupled to the member and the transition arm.
 12. The hydraulic device of claim 10 wherein the at least one compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by applying a force to the member such that a force is applied to the rotatable drum in a direction opposite to the force caused by the inlet or output pressure.
 13. The hydraulic device of claim 10 further comprising: a spring mounted between the member and the rotatable drum such that the spring applies a force to the rotatable drum in the opposite direction of the force caused by the inlet or output pressure during operation.
 14. The hydraulic device of claim 10 wherein the member comprises a support.
 15. The hydraulic device of claim 1 further comprising: a bulkhead defining: a cylinder housing the compensating piston; and a port between the cylinder housing the compensating piston and the cylinder defined by the rotatable drum to provide the fluidic communication between the cylinder defined by the rotatable drum and the compensating piston.
 16. The hydraulic device of claim 1 further comprising a bulkhead; a face valve between the bulkhead and the rotatable drum; and wherein the compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation to maintain a seal between the face valve and the bulkhead or the face valve and the rotatable drum.
 17. The hydraulic device of claim 1 further comprising: a spring coupled to the at least one compensating piston such that the spring applies a force to the rotatable drum in the opposite direction of the force caused by the inlet or output pressure during operation.
 18. The hydraulic device of claim 1 wherein the at least one compensating piston is configured to counteract the force caused by at least one of the inlet pressure and the output pressure during operation by creating a force that acts on the rotatable drum in a direction opposite to the force caused by the inlet or output pressure.
 19. The hydraulic device of claim 1 further comprising: a non-rotating member coupled to the transition arm, wherein a radial position of the non-rotating member relative to an axis of rotation of the rotatable drum is adjustable to change an angle between the transition arm and the axis of rotation of the rotatable drum.
 20. The hydraulic device of claim 19 wherein the transition arm is coupled to the piston and the non-rotating member such that a change in the angle between the transition arm and the axis of rotation of the rotatable drum changes a stroke of the piston.
 21. The hydraulic device of claim 19 wherein the non-rotating member includes teeth that mesh with a gear such that rotation of the gear adjusts the radial position of the non-rotating member relative to the axis of rotation of the rotatable drum.
 22. The hydraulic device of claim 1 further comprising: a piston joint assembly coupling the transition arm to the piston, the piston joint assembly configured to reduce centrifugal forces acting on the piston as a result of rotation of the rotatable drum and piston.
 23. The hydraulic device of claim 22 wherein the piston joint assembly comprises: a casing that defines a hole through which a portion of the piston extends, the hole having a diameter larger than the portion of the piston extending through the hole.
 24. The hydraulic device of claim 23 wherein the portion extending through the hole of the casing terminates in a head piece, the head piece having a diameter larger than the hole defined by the casing, the casing being dimension to provide a spacing between an inner surface of the casing and the head piece.
 25. A hydraulic device comprising a rotatable drum defining at least one cylinder; at least one piston housed in the cylinder; a transition arm coupled to the at least one piston and configured to translate between linear motion of the piston and rotation of the rotatable drum; and means for counteract a force caused by at least one of an inlet pressure and an output pressure during operation.
 26. A method for counteracting a force in a hydraulic device caused by at least one of an inlet pressure and an outlet pressure during operation, the method comprising: adjusting a counteracting force as at least one of the inlet pressure or the outlet pressure changes during operation.
 27. A method for counteracting a force in a hydraulic device caused by at least one of an inlet pressure and an outlet pressure, the method comprising: rotating a drum that defines a cylinder; reciprocating a piston in the cylinder; communicating a pressure in the cylinder to a compensating piston such that the compensating piston creates a force that counteracts the force caused by at least one of an inlet pressure and an outlet pressure during operation. 